Methods of supercharging a diesel engine, in supercharged diesel engines, and in supercharging units for diesel engines

ABSTRACT

An internal combustion engine of the expansible chamber type and preferably a diesel engine is equipped with a turbo-compressor unit, comprising at least one compressor and at least one turbine, and at least one bypass pipe enabling direct and permanent passage for the air delivered through the compressor to the turbine inlet. 
     The diesel engine is supercharged by the compressor driven by the turbine. Regulating means are provided to limit the rotary speed of the supercharging unit so that it operates at or above a minimum threshold value such that the engine, which has a compression ratio of less than 12, can be started and kept running at low power without difficulty. For engines requiring scavenging, throttle means with variable passage cross section are arranged so as to be traversed by generating between the upstream and downstream parts of the bypass pipe a pressure difference which is an increasing function of the pressure existing in the upstream part regardless of the engine speed and therefore which is independent of the air flow passing through said throttle means. The increasing function may be a linear or substantially linear function. A combustion chamber is arranged upstream of the turbine and supplied by air which has passed through the bypass pipe, by fuel under the control of the regulating means and also, in some embodiments, by exhaust gases from the engine.

This application is a division of my prior copending application Ser.No. 721,576 filed in the U.S. Patent and Trademark Office on Sept. 8,1976, now U.S. Pat. No. 4,125,999, which in turn is a division of myprior copending application Ser. No. 437,748, filed in the United StatesPatent and Trademark Office on Jan. 29, 1974, now U.S. Pat. No.3,988,894, which in turn is a continuation-in-part of my prior copendingapplication Ser. No. 384,566, filed Aug. 1, 1973, now abandoned, whichin turn is a continuation of my prior copending application Ser. No.139,080, filed Apr. 30, 1971, now abandoned, and claiming the benefit ofthe priority date of my French application Ser. No. 70/16289, filed May5, 1970, and which application Ser. No. 437,748 is also acontinuation-in-part of my prior copending application Ser. No. 345,968,filed Mar. 29, 1973, now abandoned, and claiming the benefit of thepriority dates of my French applications Ser. No. 72/12113, filed Apr.6, 1972, and Ser. No. 73/10041, filed Mar. 21, 1973.

The invention relates to supercharging methods for internal combustionengines of the expansible combustion chamber type, preferably a dieselengine with a supercharging unit comprising a compressor supplying freshair in parallel to the engine and to a combustion chamber, and a turbinesupplied with combustion gas by the engine and said combustion chamber,the said turbine driving in rotation the said compressor, independentstarting means being provided to bring the turbine-compressor assemblyto self-maintaining operation independent of the engine.

The invention also relates to diesel engines supercharged by asupercharging unit comprising a compressor, supplying fresh air inparallel to the engine and to a combustion chamber, and a turbinesupplied with combustion gas by the engine and the abovesaid combustionchamber, the abovesaid turbine driving said compressor in rotation,independent starting means being provided to bring theturbine-compressor assembly to self-maintaining operation independent ofthe engine.

The invention relates also to supercharging units for internalcombustion engines, preferably of the diesel type, comprising acompressor supplying fresh air to an auxiliary combustion chamber andalso to the engine combustion chamber via first connecting meansconnected to an intake manifold of the engine, and a turbine suppliedwith combustion gases provided by said auxiliary combustion chamber andby the engine combustion chamber via second connecting means connectedto the exhaust manifold of said engine, said turbine rotating saidcompressor, independent starting means being provided to bring theturbine-compressor assembly to self-maintaining operation.

As set forth in more detail following the detailed description of theembodiments shown in FIGS. 1-13, it has been appreciated that it wouldbe advantageous, especially from the point of view of specificpower/stroke volume ratio and from the point of view of robustness andsimplicity, to provide supercharged diesel engines having a lowcompression ratio, less than 12, and which can be as low a value as 8 oreven 6.

Now it is known that, all things being otherwise equal, reduction of thecompression ratio of a supercharged diesel engine causes the appearance,below a certain limiting compression ratio, which is situated around 12,of impossibilities of starting and difficulties of low power operation,and this by reason of the fact that the temperature of self-ignition ofthe fuel is no longer reached at the end of the compression stroke.

It is a specific object of the invention to improve the operation ofsupercharged diesel engines and to permit, for engines whose compressionratio is less than 12, starting without having to resort to any specialstarting method, and correct operation at idle and low power.

The supercharging method according to a preferred embodiment of theinvention is characterized by the fact that, the engine having acompression ratio less than 12, the minimal rotary speed of itssupercharging unit is limited to a threshold value sufficient to create,in the intake pipe of the engine, conditions of temperature and ofpressure enabling its starting and its operation at low power, thisthreshold value being all the higher, for a given supercharging unit, asthe compression ratio of the engine is lower.

Preferably, the abovesaid threshold value is obtained by limiting, i.e.,controlling or regulating, the supply of fuel to the combustion chamber.

The diesel engine according to a preferred embodiment of the inventionis characterized by the fact that it has a compression ratio less than12, by the fact that regulating means for the speed of its superchargingunit are provided and are arranged so that the minimal rotary speed ofthis supercharging unit is limited to a threshold value sufficient tocreate, in the intake pipe of the engine, conditions of temperature andof pressure enabling its starting and its operation at slow speed, thisthreshold value being all the higher, for a given supercharging unit, asthe compression ratio of the engine is lower.

Preferably, the abovesaid regulating means are constituted by a supplydevice limiting, i.e., controlling or regulating, the flow rate of fuelintroduced into the combustion chamber.

The supercharging unit according to a preferred embodiment of theinvention is characterized by the fact that first connecting means areprovided for the compressor to supply with fresh air, in parallel, thecombustion chamber and the intake pipe of a diesel engine with acompression ratio less than 12, by the fact that second connecting meansare provided so that the turbine can be supplied with the combustion gasby the combustion chamber and by the exhaust pipe of the abovesaiddiesel engine, and by the fact that regulating means of the speed of thesupercharging unit are provided and are arranged so that the minimalrotary speed of this supercharging unit is limited to a threshold valuesufficient to create, in the intake pipe of the engine, conditions oftemperature and of pressure enabling its starting and its operation atlow power, this threshold value being all the higher, for a givensupercharging unit, as the compression ratio of the engine concerned islower.

The abovesaid regulating means can comprise an actuating member whichcan modify the threshold value according to the compression ratio of theengine concerned.

To a first approximation, the power of an engine is proportional to theamount of air inspired. The power of a given engine, whose rotary speedis fixed, hence can only be increased at the cost of an increase in thedensity of the intake air. So it is necessary to increase the pressureand to reduce the temperature of this air. On the other hand to respectthe longevity of the engine, the maximal admissible pressure must not beexceeded and the temperature of the gases in the cylinder must not beunduly raised. A considerable increase in intake pressure is hence onlypossible at the cost of a correlated lowering of the compression ratiowhich is accompanied by lesser heating of the air during the compressionstroke. Below a limiting volumetric compression ratio of the enginecomprised between 12 and 17 according to the bore, this heating maybecome insufficient to enable self-ignition of the fuel at least atstarting, idling or lower power operation.

In accordance with one feature of the present invention, such relativelylow compression ratios comprised between about 6 and 10, as the case maybe, are employed, while around the engine an artificial atmosphere ismaintained under sufficient pressure and temperature to palliate thelack of compression in the cylinder and this pressurized atmosphere isestablished prior to the starting of the engine. The pressure of thisatmosphere is sufficiently high to enable easy starting up at the lowestambient temperatures. The low compression ratio gives the engine asmoothness of operation which is manifested by a considerable reductionin the characteristic knock or chatter of a diesel engine and less wearof the moving connecting parts. It makes possible an excess of air inthe cylinder, which in turn lowers the maximal and average temperaturesof the gases and hence the thermal load on the engine. Moreover, thisexcess of air reduces the creation of nitrogen oxide (due to thelowering of temperatures) and the formation of unburnt products and ofsmoke (due to the excess of oxygen at all speeds).

To provide the high pressure of air required, the turbocompressor isoperated within the narrow range of good compression yields, that is tosay like a gas turbine. The ability to operate in this high yield modeis achieved by the system of the invention due to the previouslydescribed parallel connection in the form of a bypass connecting thecompressor outlet to the turbine inlet whose permeability is controlledto maintain the good yield of the compressor. This bypass also suppliesfresh air to a combustion chamber which is situated upstream of theturbine and which enables self-sustaining operation of the turboblower.The latter can then be started up prior to the engine and kept above aminimal speed in the whole range of operation of the engine. Theartificial pressurized atmosphere defined above is thus realized.

The above features enable, moreover, the obtaining of maximum torque atall rotary speeds, the supercharging pressure being adjustableindependently of the engine speed. The increase in power is not obtainedby increasing the forces on the connecting rod system but by increasingthe duration during which they are applied. Therefore, the basicconstruction of the engine can be preserved in adapting a conventionalengine to the system of the invention. Similarly, since the averagetemperature of the gases in the cylinder is lower, the water circulationof the original engine is sufficient for the increase in power and hencethe original cooling system need not be modified.

The method of the invention is applicable to any existing self-ignitioninternal combustion engine and enables the obtaining of three to fourtimes the power of the unsuper-charged engine without changing itsoperating life. Relative to conventional supercharging, the power gainthus obtained can go from 50% to 150%, depending upon the ratio ofinitial supercharging employed. The increase in power is alsoaccompanied by a certain number of secondary advantages: high torque atlow speed, much reduced noise level, exhaust of low polluting effect,ease of cold starting, easy correction of atmospheric variations,possibility of idling at very low speed, and reduction in the specificbulk of the cooling system.

The system of the invention is readily adapted to the majority ofself-ignition internal combustion engines. It requires no internalmodification other than a different geometry of the combustion chambers.Moreover, the very high pressure necessary for the method is provided bya supercharging system which is used instead and in place ofconventional supercharging devices of similar bulk.

In such engines of the invention, the parallel branch of the aforesaidconnecting means preferably comprises a bypass pipe enabling direct andpermanent passage of fresh air delivered by the compressor to theexhaust gases emerging from the engine. A combustion chamber is thengenerally provided upstream of the turbine, this combustion chamberbeing supplied by the exhaust gases and by the fresh air taken from theabovesaid branch pipe.

It is a further object of the invention to adapt the turbo-compressorgroup to high supercharging pressures due to operation of the compressorclose to the surge or pumping line, that is to say with optimum yield.

It is yet another object of the invention to enable good scavenging ofthe engine due to a difference of pressure maintained between the intakeand the exhaust.

It is another object of the invention, in engines of the abovecharacter, to reduce the work of discharging exhaust gases, which henceenables the power of the engine to be increased (by increasing the meaneffective pressure) and to reduce its consumption.

In order to achieve such scavenging the engine according to anotherembodiment of the invention is provided with throttle means withvariable passage cross section, arranged so as to be traversed by theair passing through the bypass pipe, these throttle means generatingbetween the upstream part of the bypass pipe (the part connected to thecompressor) and the downstream part of the bypass pipe (the partconnected to the turbine, if necessary through the combustion chamber) adifference in pressure which is an increasing function, preferablylinear or substantially linear, of the pressure existing in the upstreampart regardless of the engine speed and therefore which is independentof the air flow passing through said throttle means.

It will hence be understood that the work of discharging the exhaustgases being reduced, the brake mean effective pressure (b.m.e.p.) isincreased to a value equal to the aforesaid difference in pressurebetween the pressure upstream of the throttle means and the pressuredownstream of said throttle means.

Moreover, it is possible to make the engine operate at highsupercharging pressures, the compressor operating close to the pumpinglimit.

Lastly, the existence of a difference in pressure maintained between theintake (pressure upstream of the throttle means) and the exhaust(pressure downstream of the throttle means) enables good scavenging ofthe engine.

According to one advantageous embodiment of the invention, the throttlemeans comprises a throttle member arranged in the bypass pipe andcooperating with a fixed seat.

This throttle member can be operatively coupled to, or may consist of, abalancing piston, one working surface of which is subjected to thepressure existing in the part upstream of the bypass pipe and of which asecond working surface is subject to a counter pressure (atmosphericpressure or pressure comprised between atmospheric pressure and thepressure existing in the upstream part of the bypass pipe), and a thirdworking surface of which is subjected to the pressure existing in thedownstream part, and elastic return means being able to act in one senseor the other on the movable mechanism constituted by the throttle memberand its balancing piston.

According to a particular feature of the invention, which is applied inthe case where there is provided a combustion chamber which is supplied,with fresh air, through a primary air intake to introduce fresh air intoa combustion zone, and through a secondary air intake to introduce freshair into a mixing zone, the throttle means comprise, in parallel, firstthrottle means with variable passage cross section, arranged so as to betraversed by the secondary air, these first throttle means generatingbetween the upstream part of the bypass pipe (the part connected to thecompressor) and the downstream part of the bypass pipe (the partconnected to the combustion chamber) a pressure difference which is anincreasing function, preferably linear or substantially linear, of thepressure existing in the upstream part, and second throttle means withvariable outlet cross section subjected to the difference of pressuregenerated by the first throttle means and arranged so as to be traversedby the primary air, these second throttle means regulating the flow-rateof primary air by offering an outlet cross section to this primary airwhich is servocoupled to the pressure existing in the downstream part orthe upstream part of the bypass pipe, this servocoupling being accordingto a predetermined relationship.

Preferably, these second throttle means control in addition a regulatingdevice for the flow-rate of fuel injected into the combustion chamber soas to preserve, for flow-rates of primary air and of fuel, a ratioensuring good combustion stability.

The invention, apart from the features which have been considered,consists of certain other objects, features and advantages which will bemore explicitly discussed below.

The invention will, in any case, be better understood with the aid ofthe supplementary description which follows as well as of theaccompanying drawings, which description and drawings relate topreferred embodiments of the invention and do not have, of course, anylimiting character. In these drawings:

FIG. 1 is a diagrammatic view of a supercharged diesel engine, with onecompression stage, constructed according to the invention andconstituting a first embodiment.

FIG. 2 is a diagrammatic view of a supercharged diesel engine, with onecompression stage, constructed according to the invention andconstituting a second embodiment thereof.

FIG. 3 is a diagrammatic view of a supercharged diesel engine, with onecompression stage, constructed according to the invention andconstituting a third embodiment thereof.

FIG. 4 shows a supercharging unit according to the invention, with onecompression stage, and constructed in a manner analogous to theembodiment of the diesel engine shown in FIG. 1.

FIG. 5 shows a supercharging unit according to the invention, with onecompression stage, and constructed according to the invention in amanner analogous to the embodiment of the diesel engine shown in FIG. 2.

FIG. 6 shows a supercharging unit according to the invention, with onecompression stage, and constructed in a manner similar to that of theembodiment of the diesel engine shown in FIG. 3.

FIGS. 7 and 8 show two variations of a supercharged diesel engine withtwo compression stages and constructed according to the invention in amanner similar to the embodiment of the supercharged diesel engine withone compression stage which is illustrated in FIG. 2.

FIG. 9 is a diagrammatic view of another embodiment of a superchargeddiesel engine equipped with a combustion chamber with a single fresh airintake, and constructed according to the invention.

FIGS. 9A and 9B are diagrammatic views of second and third embodimentsof throttle means which are equivalent to and may be substituted for thethrottle means shown in FIG. 9.

FIG. 10 is a diagrammatic view of a supercharged diesel engine, equippedwith a combustion chamber with two fresh air intakes, and constructedaccording to an embodiment of the invention for which the combustionchamber comprises a "return" injector.

FIG. 11 is a diagrammatic view of a supercharged diesel engine, equippedwith a combustion chamber with two fresh air intakes, and constructedaccording to an embodiment of the invention for which the combustionchamber comprises a "nonreturn" injector.

FIG. 12 is a partial view of an important element of the engine of FIG.10 showing a modification of the invention.

FIG. 13 is a graph relating to the operation of an engine according tothe invention.

FIG. 14 is a schematic diagram of a commercial Poyaud Model 520-6Lengine modified in accordance with the present invention.

FIG. 15 is a rear perspective elevational view of a commercial PoyaudModel 520-6L engine also modified in accordance with the presentinvention and in particular utilizing the system shown in FIG. 2.

FIG. 16 is a front perspective elevational view of the commercial PoyaudModel 520-6L engine incorporating the systems described in conjunctionwith FIGS. 11-14.

FIG. 17 is a graph showing a plot of engine r.p.m. against horsepowerand against specific fuel consumption for nonsupercharged,conventionally supercharged and invention supercharged versions of theengine shown in FIGS. 14-16.

FIG. 18 is a graphic comparative development of supercharging pressureversus crank angle for engines supercharged conventionally and pursuantto the invention, and FIGS. 19, 20 and 21 are corresponding graphiccomparative developments of temperature, heat transfer coefficients andthermal flow respectively.

FIG. 22 is a graph of indicated mean effective pressure as a function ofthe supercharging pressure and of the angular duration of combustioncalculated for an engine supercharged pursuant to the invention.

FIG. 23 is a graph of the paths of operative points of a turbocompressorin its characteristic field as operated in conventional superchargingand in supercharging pursuant to the invention.

FIG. 24 is a graphic illustration of three phases of operation of engineoperated pursuant to the method of the invention depicted as plot ofhorsepower versus engine r.p.m.

As shown in FIGS. 1-3, the diesel engine 1 is supercharged by asupercharging unit, with one compression stage, which comprises acompressor 2, supplying fresh air in parallel to the engine 1 and to acombustion chamber 3, and a turbine 4 supplied with combustion gas bythe engine 1 and by the abovesaid combustion chamber 3. The turbine 4rotates the compressor 2 through a connecting shaft 5. Independentstarting means 6, which can be constituted by an electric motorassociated with a clutch, are provided to bring the turbine 4-compressor2 assembly into self-maintaining operation independent of the engine.

This engine has a compression ratio less than 12, and regulating means,that is to say, threshold fixing means for regulating the speed of thesupercharging unit are provided and are arranged so that the minimalrotary speed of this supercharging unit is limited to, i.e., does notdrop below, a threshold value sufficient to create, in the intake pipe1a of the engine 1, conditions of temperature and of pressure enablingits starting and its operation at slow speed.

This threshold value is all the higher, for a given supercharging unit,as the compression ratio of the engine is lower.

The said regulating means are constituted by a supply device 7 limitingor controlling the flow rate of fuel introduced into the combustionchamber 3, and such a supply device can be constituted by anelectro-pump delivering a flow rate of fuel proportional to its supplyvoltage.

In the embodiment illustrated in FIG. 1, the combustion chamber 3receives only fresh air coming from the compressor 2.

In the embodiment illustrated in FIG. 2, the combustion chamber 3receives, on one hand, fresh air coming from the compressor 2, and, onthe other hand, all or part of the combustion gases coming from theexhaust pipe 1b of the engine 1.

In the embodiment illustrated in FIG. 3, the combustion chamber 3receives a mixture constituted by fresh air coming from the compressor 2and by all or part of the combustion gases coming from the exhaust pipe1b of the engine 1.

By way of example, if it is assumed that the diesel engine has acompression ratio of 8 and that the turbine and the compressor of thesupercharging unit have an isentropic yield (adiabatic efficiency) of0.75, there will be obtained for this supercharging unit aself-maintaining operation for a compression ratio in the compressor of1.3, which will produce a rise in temperature in the compressor of about30° C.

Now, in an engine of compression ratio 8, the minimal temperature in theintake pipe is about 80° C. to obtain the auto-ignition temperature ofthe fuel at the end of the compression stroke.

To obtain this value of 80° C. for an ambient temperature which will beassumed equal to 0° C., it is hence necessary that the temperature ratioin the compressor be equal to 1.3, or with an insentropic yield of 0.75,a pressure ratio equal to 2 in the compressor. It is this pressure ratiowhich fixes the threshold value of the rotary speed of the superchargingunit.

There will now be described, with reference to FIGS. 4-6, asupercharging unit with one compression stage, for a diesel engine. Thissupercharging unit comprises, a compressor 20 supplying fresh air to acombustion chamber 30, and a turbine 40 supplied with combustion gas bythe abovesaid combustion chamber 30. The turbine 40 rotates thecompressor 20 through a connecting shaft 50. Independent starting means60, such as for example an electric motor associated with a clutch, areprovided to bring the turbine 40-compressor 20 assembly toself-maintaining operation.

First connecting means 21 are provided in order that the compressor 20may supply fresh air, in parallel to the combustion chamber 30 and theintake pipe of a diesel engine (not shown) with a compression ratio lessthan 12. Second connecting means 41 are provided so that the turbine 40can be supplied with combustion gas by the combustion chamber 30 and bythe exhaust pipe of the abovesaid diesel engine.

Regulating means for the speed of the supercharging unit are providedand are arranged so that the minimal rotary speed of this superchargingunit is limited to a threshold value sufficient to create, in the intakepipe of the engine, conditions of temperature and of pressure enablingits starting and its operation at slow speed. This threshold value is,all other conditions being kept the same, all the higher, for a givensupercharging unit, as the compression ratio of the corresponding engineis low. These regulating means can be constituted by a supply device 70for controlling the flow rate of fuel introduced into the combustionchamber 30; such a supply device can be constituted by an electro-pumpdelivering a flow rate of fuel proportional to its supply voltage. Theseresulting means can additionally comprise an actuating member which canmodify the threshold value according to the compression ratio of theengine concerned.

In the embodiment illustrated in FIG. 4, the combustion chamber 30 onlyreceives fresh air coming from the compressor 20 when the superchargingunit cooperates with the engine.

In the embodiment illustrated in FIG. 5, the combustion chamber 30receives, when the supercharging unit cooperates with the engine, on onehand, fresh air coming from the compressor 20, and, on the other hand,all or part of the combustion gases coming from the exhaust pipe of thesaid engine.

In the embodiment illustrated in FIG. 6, the combustion chamber 30receives, when the supercharging unit cooperates with the engine, amixture constituted by fresh air coming from the compressor 20, and byall or part of the combustion gases coming from the exhaust pipe of theabovesaid engine.

In the foregoing, it has been assumed that the supercharged dieselengine according to the invention included a supercharging unit with onecompression stage, but such a diesel engine can include a superchargingunit with two or more compression stages.

A supercharged diesel engine comprising a supercharging unit with twocompression stages is illustrated in FIGS. 7 and 8, this diesel enginebeing assumed constructed in a manner similar to that of the embodimentillustrated in FIG. 2, the same reference numerals denoting the samemembers in FIGS. 7 and 8 and in FIG. 2, but modified by the index 1 whenthey relate to elements of the high pressure stage. This diesel enginehence comprises a supercharging unit with a low pressure stage,compressor 2-turbine 4, and a high pressure stage, compressor 2₁-turbine 4₁.

In the variation illustrated in FIG. 7, it is the high pressurecompressor 2₁ which supplies the combustion chamber 3, whilst in thevariation illustrated in FIG. 8 it is the low pressure compressor 2which supplies the combustion chamber 3. In FIG. 7 starting means 6₁ arecoupled to turbocompressor 2₁ -4₁ whereas in FIG. 8 starting means 6 arecoupled to turbocompressor 2-4.

Of course, there can be envisaged other supercharging arrangements withtwo or several compression stages and the invention can be supplied alsothereto.

It is the same for supercharging units which could have also two orseveral compression stages arranged in different ways, the inventionbeing then applicable to such supercharging units.

The method of supercharging a diesel engine according to the inventionis thus also applicable to a diesel engine with a supercharging unitwith two or several compression stages arranged according to varioussystems.

The invention also enables the exploitation of supercharged dieselengines of which the compression ratio is less than 12, these dieselengines being capable of being started without resorting to any specialstarting process and capable of operating correctly at slow speed.

Moreover, the invention enables the obtaining, for such diesel enginesof a satisfactory overall yield (said yield taking into account amountsof fuel injected into the diesel engine and into the combustionchamber).

Referring now in more detail to the embodiment shown in FIG. 9, thediesel engine shown therein is denoted by the reference numeral 101 andit is supercharged by a turbo-compressor unit denoted by the referencenumeral 102.

This turbo-compressor unit 102 comprises a compressor 103 deliveringcompressed air to supply the engine through a pipe, and a turbine 104driving said compressor 103 through a shaft 105, this turbine 104 beingactuated by exhaust gases from the engine 101.

There is provided a bypass pipe 106 permitting direct and permanentpassage of fresh air taken from the compressor 103 to the turbine 104.

Preferably a combustion chamber 107 is then provided upstream of theturbine 104, this combustion chamber 107 being supplied with exhaustgases from the internal combustion chamber of engine 101 and with freshair taken from the bypass pipe 106.

According to this embodiment of the invention, there is providedthrottle means 108 with variable outlet cross section, arranged so as tobe traversed by the air flowing in the bypass pipe 106. As will beexplained in more detail hereinafter, these means 108 generate betweenthe upstream part of the bypass pipe 106 (the part connected to thecompressor 103) and the downstream part of the bypass pipe 106 (the partconnected to the turbine 104 through the combustion chamber 107) adifference of pressure ΔP which is an increasing function, preferablylinear or substantially linear, of the pressure P existing in theupstream part, and hence which pressure difference ΔP is independent ofthe air flow rate through bypass pipe 106. This linear function may bewritten:

    ΔP=α'P+β'

α' and β' denoting two coefficients.

In the embodiment of the invention illustrated in FIG. 9, these throttlemeans 108 comprise a throttle member 108a arranged in the bypass pipe106 and cooperating with a fixed seat 108e.

This throttle member 108a can be carried by a stem 108b which slides ina cylinder, or better, which is connected to the bypass pipe 106 througha deformable wall 108d. The relationship between the diameter of thethrottle member 108a and the diameter of the piston 108c establishes thecoefficient a' and is such that the said throttle member 108a isbalanced (as set forth in more detail hereinafter), by the pressure Pexerted on its upstream surface and the inner surface of the piston108c, by the pressure P-ΔP exerted on its downstream surface, and by theatmospheric pressure exerted on the outer surface of the piston 108c orby a suitable counter pressure source.

Elastic return means may also act on the throttle member 108a. Thiselastic return means and the aforesaid counter pressure fix the value ofthe coefficient β' of the relationship ΔP=α'P+β'. These elastic returnmeans may be constituted by a spring 109 and/or by the elasticity itselfof the deformable wall 108d.

To permit the adjustment of this coefficient β', there may be providedregulating means enabling adjustment of the resultant force applied onthe throttle member 108a by the elastic return means; these regulatingmeans can be constituted by a nut 110 modifying the tension of theelastic return means, said but being carried by the outer threaded partof stem 108b and abutting piston 108c.

This feature is particularly advantageous since it enables the pressuredifference created by the throttle means 108 to be adapted to theinstallation of which the supercharged engine forms part. In particular,this pressure difference can be adapted to the pressure losses which arecreated by a filtering device 111 situated at the intake of thecompressor 103 and/or by a silencer device 112 located at the outlet ofthe turbine 104.

A viscous damping device, which will be more explicitly consideredbelow, can advantageously act on the throttle member 108a in order toabsorb vibration of aerodynamic origin to which said throttle member108a can be subjected. This viscous damping device is preferablysupplied from a source of viscous fluid under variable pressure.

Referring to FIGS. 9A and 9B, there are diagrammatically illustratedalternative throttle means which may be substituted for the abovedescribed throttle means 108-108e, 109, 110 in the bypass pipe 106. Suchalternative throttle means are also disclosed in greater detail and areclaimed in a copending joint application of the inventor herein, Jean F.Melchior, and Thierry Andre, Ser. No. 615,775, filed Sept. 22, 1975, andassigned to the assignee herein. In FIGS. 9A and 9B a movable throttlemember 120 is connected via a linkage 122 to a balancing piston 124which is made up of a large diameter piston 126 connected by a rod 128to a small diameter piston 130. Piston 126 slides in a cylinder 132 andpiston 130 in a cylinder 134. A compression coil spring 136 is receivedin cylinder 134 to bias piston 124 to the right as viewed in FIGS. 9Aand 9B. A conduit 138 provides communication between the portion of pipe106 downstream of throttle 120 and chamber 140 of cylinder 132.

In FIG. 9A another conduit 142 connects the portion of pipe 106 upstreamof throttle 120 with a chamber 144 of cylinder 132. The spring chamber146 of cylinder 134 is connected to a source of the aforesaid counterpressure P_(o). It will be observed that movement of piston 124 to theright opens throttle 120 and vice versa. Piston 126 has oppositelyfacing working surfaces of area S, and piston 130 has a working surfaceexposed to chamber 146 of area s. The values of S, s, counter pressureP_(o) and the force F of spring 136 are suitably chosen such that thepressure drop ΔP equals the expression: ##EQU1## for the system of FIG.9A, and this relationship will hold true regardless of the value of Qwhich is the air flow-rate passing through bypass 106 and past throttle120. Thus, an increase in air flow rate Q for a given pressure P willproduce an increase in the pressure drop ΔP downstream of throttle 120.This causes a decrease in the pressure in chamber 140 causing piston 126to move to the right, thereby opening throttle 120 so as to reduce theamount of the pressure drop ΔP until it returns to the initial valuethereof to re-establish the initial ratio of the pressure drop ΔP to##EQU2## Likewise, a decrease in Q will produce an opposite motion inthrottle 120 to maintain that fixed ratio of ΔP to the value: ##EQU3##

In the system of FIG. 9B the pressure drop ΔP is equal to theexpression: ##EQU4## Likewise, the system functions to maintain aconstant ratio between ΔP and a value equal to: ##EQU5## regardless ofthe value of Q. It will be noted that in FIG. 9B the upstream pressure Pis communicated via conduit 148 to chamber 146 instead of to chamber144, and the counter pressure P_(o) is connected to chamber 144 insteadof to chamber 146.

There will now be described more particularly a feature of the inventionwhich is applied in the case where the combustion chamber is suppliedwith fresh air, through a primary air intake to introduce fresh air intoa combustion zone, and through a secondary air intake to introduce freshair into a mixing zone.

By way of example, this feature may be accomplished by first throttlemeans which comprise a movable mechanism which houses, on one hand, thesecond throttle means and, on the other hand, the regulating device forthe flow-rate of fuel.

To this end, the movable mechanism of the first throttle means can beconstituted by a cylinder bearing on the outside a throttle membercooperating with a fixed seat, the second throttle means being thenconstituted by one or several ports formed in this cylinder and by aslide valve covering or uncovering this or these ports, this slide valvebeing actuated by a piston of which one of the faces is subjected to thepressure existing in the downstream or upstream part of the bypass pipe,and the other face to the action of a counter pressure and to the actionof a spring, this slide valve or this piston being advantageouslyconnected to the regulating device for the flow-rate of fuel.

In this respect, reference will be made first of all to FIGS. 10 and 11,which show a diesel engine denoted by the reference numeral 201; thisengine is supercharged by a turbo-compressor unit denoted by thereference numeral 202.

This turbo-compressor unit 202 comprises a compressor 203 deliveringcompressed air to supply the engine through a pipe 204, and a turbine205 driving said compressor 203 by means of a shaft 206, this turbine205 being actuated by the exhaust gases from the engine 201.

There is provided a bypass pipe 207 permitting direct and permanentpassage of fresh air delivered by the compressor 203 to the exhaustgases emerging from the engine.

A combustion chamber 208 is then provided upstream of the turbine 206,this combustion chamber 208 being supplied by the exhaust gases througha pipe 209, and by fresh air taken from the bypass pipe 207. Thiscombustion chamber 208 is in addition supplied with fuel by means of aninjector 210 or 210a supplied by a pump 211 or 211a from a reservoir212.

As regards the supply of this combustion chamber 208 with exhaust gasesand with fresh air, recourse is had to an arrangement according to whichsaid combustion chamber is supplied, by a primary air intake 213 tointroduce fresh air into a combustion zone 214, by an exhaust gas intake215 to introduce these exhaust gases into a mixing zone 216 situateddownstream of the combustion zone 214, and by a secondary air intake 217to introduce fresh air at the level of the abovesaid mixing zone 216.

The primary air intake 213 can be constituted by a central pipe 218arranged coaxially with the combustion chamber 208. The exhaust gasintake 215 can then be constituted by a first annular pipe 219encircling the central pipe 218. Lastly, the secondary air intake 217can be constituted by a second annular pipe 220 encircling the firstannular pipe 219.

According to the invention, there is provided first throttle means 221with variable outlet cross section arranged so as to be traversed by thesecondary air, these first throttle means 221 generating between theupstream part of the bypass pipe 207 (part connected to the compressor203) and the downstream part of the bypass pipe 207 (part connected tocombustion chamber 208); a pressure differential ΔP which is anincreasing function, preferably linear or substantially linear, of thepressure P existing in the upstream part regardless of the engine speedand, therefore, which is independent of the air flow passing throughthrottle means 221. A second throttle means 222 is also provided withvariable outlet cross section subjected to the pressure difference ΔPgenerated by the first throttle means 221 and arranged so as to betraversed by the primary air, these second throttle means 222 offeringan outlet cross section to this primary air which is servocoupled to thepressure P-ΔP existing in the downstream part of the bypass pipe 207,this servocoupling being according to a predetermined law.

Advantageously, these second throttle means 222 control in addition aregulating device for the flow rate of fuel 223 controlling the amountof fuel injected into the combustion chamber 208 so as to preserve, forflow rates of primary air and of fuel, a ratio insuring good stabilityof combustion, that is to say, a relationship as close as possible tostoechiometric proportions.

It will then be understood that the fresh air taken from the bypass pipe207 is divided into primary air coming from the primary air intake 213and the second throttle means 222 to supply the combustion zone 214 ofthe combustion chamber 208 and into secondary air coming from thesecondary air intake 217 and the first throttle means 221 to supply themixing zone 216 of the combustion chamber 208. The flow rate of aircirculating in the bypass pipe 207 varies in a ratio of the order of 1to 10 according as the engine operates at full speed (combustion chamberturned low) or as the engine operates at idling speed (combustionchamber used to the maximum).

By way of explanation of the aforementioned first and second throttlemeans associated with such a combustion chamber, it will then beunderstood that there is established a relationship between the outletcross section S_(p) offered to the primary air and the outlet crosssection S_(s) offered to the secondary air. In fact, if ΔP denotes thepressure difference on each side of the first throttle means 221 and Pthe pressure existing in the upstream part of the bypass pipe, theincreasing linear function connecting ΔP and P can be written in thefollowing way: ΔP=αP+β, α and β denoting two coefficients. On the otherhand, it can be written that this pressure difference ΔP is proportionalto the specific mass m of the fresh air and to the square of itsvelocity V: ΔP=km V², k being a constant to a first approximation. Fromthe two above equations, there can then be deducted the value of thevelocity V: ##EQU6## Now the sum of the outlet cross sections S_(p) andS_(s) is connected to the total flow rate Q of fresh air in the bypasspipe by the following equation: ##EQU7## with S_(p) being a function ofP, namely S_(p) =f(P), f(P) being the predetermined relationship betweenthe outlet cross section S_(p) and the pressure P, namely, by replacingthe velocity by its value as a function of the pressure: ##EQU8## Underthese conditions, the flow rate of primary air Q_(p) only depends on thepressure P according to the equation: As will be evident, the flow rateof secondary air Q_(s) is always equal to the difference between thetotal flow rate flowing in the bypass pipe and the flow rate of primaryair Q_(p) : Q_(s) =Q-Q_(p). On the other hand, the flow rate of fuelwhich is necessary to insure the operation of the combustion chamber 208enabling autonomy of the turbo-compressor unit 202 on starting of theengine and the flow rate of fuel sufficient to maintain the operation ofthe combustion chamber 208 turned low are within a ratio of the order of30 to 1.

Now the stability of combustion in the combustion chamber will beoptimal if the flow rate of air insuring the combustion, that is, theflow rate of air coming from the primary air intake, is in theneighborhood of the flow rate corresponding to stoechiometricproportions.

As previously explained, the sum of the outlet cross sections S_(p)+S_(s) (respectively offered to the primary and to the secondary air) ishence determined by the values of the delivery pressure P from thecompressor and of the flow rate Q of fresh air in the bypass pipe 207.Under these conditions and to organize the flow rates of primary andsecondary air, it suffices to act on one of these outlet cross sections,the second being adjusted itself by the difference. Action is thereforeon the outlet cross section S_(p) offered to the primary air.

As regards the predetermined relationship followed by the servocouplingbetween, on one hand, the outlet cross section S_(p) offered to theprimary air and, on the other hand, the delivery pressure P-ΔP, whichexists in the downstream part of the bypass pipe 207, it can be selectedto respect the operation of the combustion chamber with theengine-turbo-compressor group as a whole. This law will be consideredagain in more detail below with respect to the two embodiments of theinvention relating, the first, to a supply of the combustion chamber bya "return" type injector (FIG. 10) and, the other, to supply of thecombustion chamber by a "nonreturn" type injector (FIG. 11).

In a particularly simple embodiment of the invention from the point ofview of its construction, and which can provide two forms of themachine, illustrated respectively in FIGS. 10 and 11, the first throttlemeans 221 comprise a movable mechanism which houses the second throttlemeans 222 and the regulating device for the flow rate of fuel 223. Tothis end, the movable mechanism of the first throttle means 221 can beconstituted by a cylinder 224 bearing on the outside a throttle member225 cooperating with a fixed seat 226. Under these conditions, thesecond throttle means 222 can then be constituted by one or severalports 227 formed in this cylinder 224 and by a slide rod 228 covering oruncovering this or these ports 227, this slide rod 228 forming a pistonof which one of its faces is subjected to the pressure of the primaryair and the outer face to the action of a counter pressure P_(c) and tothe action of a spring 229, this slide valve 228 being connected to theregulating device for the flow rate of fuel 223. The covering anduncovering of the one or more ports 227 by the slide valve 228 forming apiston can be effected by providing said piston with one or severalorifices 228a formed in its skirt 228b.

As regards the counter pressure exerted on this slide valve 228, it canbe equal to atmospheric pressure. However, in certain cases, and toshift the area of regulation, a counter pressure P_(c) can be selectedhigher than atmospheric pressure. It is then particularly simple toprovide a regulating chamber 230 communicating with the surfaceconcerned of the slide valve 228 and communicating also with atmosphericpressure through a throttle orifice 231, this regulating chamber 230being supplied by air under pressure by means of a flexible pipe 232.This air under pressure escapes permanently from the regulating chamber230 through a throttle orifice 231 and causes in said regulating chamber230 an excess pressure which then constitutes the counter pressureP_(c).

To this end, it is particularly advantageous to connect the flexiblepipe 232 opening into this regulating chamber 230 to a pipe from themotor installation in which cooled compressed fresh air flows. There canthen be arranged between the engine 201 and its intake cooler R asampling pipe 252 ending at the flexible pipe 232 after having traverseda needle regulating device 253 enabling adjustment of the counterpressure P_(c) to a value comprised between atmospheric pressure and thedelivery pressure P.

A regulating spring 233 is provided to act on the movable mechanism ofthe first throttle means 221, this spring 233 enabling adjustment of theparameter β of the linear function ΔP=αP+β giving the pressuredifference generated by the abovesaid first throttle means 221. Theaction of this adjusting spring 233 can advantageously be adjusted byacting on a movable stop 234 against which said spring 233 is supported.However, the first throttle means 221 can be subjected to vibrations ofaerodynamic origin, and it is advantageous to subject the movablemechanism of these first throttle means 221 to the effect of a viscousdamping device 263 shown in FIG. 12 in which the same reference numeralsdenote the same members as in FIG. 10. This viscous damping device 263can be supplied from a source of viscous fluid 264 under variablepressure. It will be evident from FIG. 12 that the device 263 mayconsist of a piston having restricted orifice perforations and connectedby a rod to cylinder 224, the opposed working surfaces of this pistondiffering by the cross section of the connecting rod to provide thedifferential action. Then the pressure of this source of viscous fluidreplaces the effect of the spring 233 and its variation replaces theaction of the movable stop 234.

As regards the regulating device for the flow rate of fuel 223, recoursemay be had, for example, to two embodiments illustrated respectively inFIGS. 10 and 11. For the embodiment illustrated in FIG. 10 correspondsto an injector 210 supplying the combustion chamber 208 of the "return"type. The pump 211 supplies the injector 210 at constant pressure bymeans of a supply passage 235 and the surplus of fuel, which is notinjected into the combustion chamber, takes a return passage 236 ending,by means of a flexible pipe 237, at the regulating device for the flowrate of fuel 223. This regulating device for the flow rate of fuel 223may comprise a movable member 238 connected by a stem 269 to the slidevalve 228 of the second throttle means 222, the adjustment action beingobtained by a variable nozzle 239. This variable nozzle can have avariation of cross section of continuous type (needle more or lessengaged in an orifice) or a variation of discontinuous type (covering oruncovering of ports). Preferably, this variable nozzle 239 comprises atleast one needle 268 with a conical portion more or less engaged in anorifice 270. The fuel brought in through the flexible pipe 237 occupiesan intake chamber 240, then is obliged to pass through the variablenozzle 239 before emerging into a recovery chamber 241, whence itreemerges by means of a flexible pipe 242 to be then sent back to thereservoir 212 through a low pressure passage 243.

Referring to FIG. 10, it is noted that when the pressure which isexerted on the slide valve 228 of the second throttle means 222increases, the slide valve 228 moves toward the left which has theeffect, on one hand, of reducing the outlet cross section of the secondthrottle means 222 and through this fact reducing the flow rate ofprimary air and, on the other hand, of displacing the movable member 238with needle 268 of the device for regulating the flow rate of fuel 223toward the left and of increasing the outlet cross section of thevariable nozzle 239, hence of increasing the flow rate of fuelcirculating in the return passage 236, and hence of reducing the flowrate of fuel injected into the combustion chamber 208. Of course, whenthe pressure which is exerted on the slide valve 228 diminishes, thereverse phenomena are produced.

In FIG. 11, the same reference numerals denote the same members as inFIG. 10, but the combustion chamber 208 is supplied with fuel by meansof an injector 210a which is of the "nonreturn" type. This injector 210ais supplied by a pump 211a from the reservoir 212 through the regulatingdevice for the flow rate of fuel 223 and a supply channel 235a. Thisregulating device for the flow rate of fuel 223 can comprise a movablemember 238a connected by a rod 269a to the slide valve 228 of the secondthrottle means 222, the adjustment action being obtained by a variablenozzle 239a. This variable adjustment can present a continuous variationof cross section (needle more or less engaged in an orifice) or adiscontinuous variation of the cross section (covering or uncovering ofports). Preferably, this variable nozzle 239a comprises at least oneneedle 268a with a conical part more or less engaged in an orifice 270a.

The fuel brought in through a flexible pipe 237a occupies an intakechamber 240a, then is obliged to pass through the variable nozzle 239abefore emerging into a recovery chamber 241a whence it emerges by meansof a flexible pipe 242a to be then directed toward the injector 210athrough a high pressure passage 243a.

To maintain a constant pressure difference on each side of the variablenozzle 239a, recourse is had to a slide valve regulator 244. This slidevalve regulator 244 is constituted by a cylinder 245 in which a freepiston 246 slides covering or uncovering a port 247 formed in the wallof this cylinder 245. This free piston hence defines, on one hand, achamber 248 situated on the side of the port 247 and placed incommunication with the intake chamber 240a of the regulating device forthe flow rate of fuel 223 and, on the other hand, a chamber 249 situatedopposite the port 247 and placed in communication with the recoverychamber 241a of the regulating device for the flow rate of fuel 223.This free piston 246 is hence subjected to the difference in pressureexisting between the intake chamber 240a and the recovery chamber 241aand to the action of a spring 250 acting contrary to this difference inpressure. The port 247 communicates with a discharge pipe 254 which endsat the fuel reservoir 212.

Referring to FIG. 11, it is noted that when the pressure which isexerted on the slide valve 228 of the second throttle means 222increases, this slide valve 228 moves toward the left, which has theeffect, on one hand, of reducing the outlet cross section of the secondthrottle means 222 and through this fact reducing the flow rate ofprimary air and, on the other hand, of moving the movable member 238awith the needle 268a of the regulating device for the flow rate of fuel223 toward the left and of reducing the outlet cross section of thevariable nozzle 239a, hence of reducing the flow rate of fuel directedto the injector 210a through the supply passage 235a. Of course, whenthe pressure which is exerted on the slide valve 228 diminishes, thereverse phenomena are produced.

As for the slide valve regulator 244, its operation is such that thefree piston 246 more or less closes the port 247. The movement of thefree piston 246 remaining very slight, the force exerted by the spring250 on the said free piston 246 is practically constant; and throughthis fact, the pressure difference between the intake chamber 240a andthe recovery chamber 241a is substantially constant whatever thepressure existing in the combustion chamber 208 and whatever the flowrate of fuel injected. The value of this pressure difference is adjustedby the biasing alone of the spring 250 for which it is possible toprovide a movable stop 251 against which said spring 250 is supported.It is then understood that the flow rate of fuel through the variablenozzle 239a depends only on the outlet cross section, and hence dependsonly on the supercharging pressure through the position of the piston228 of the second throttle means 222.

When recourse is had to one or other of the two embodiments of theinvention illustrated in FIGS. 10 and 12 or in FIG. 11, by acting on thegeometry of the second throttle means 222 and on the law of the spring229 acting on the slide valve 228 of these second throttle means 222, itis possible to select the servocoupling law between the outlet crosssection S_(P) and the downstream pressure P-ΔP, or the differencebetween this downstream pressure P-ΔP and the counter pressure P_(c)existing in the regulation chamber 230. To each value of the outletcross section S_(P) corresponds a value of the flow rate of primary airQ_(P), hence a value of the flow rate of fuel Q to be introduced intothe combustion chamber, this flow rate of fuel being insured by thegeometry of the variable nozzle.

Taking into account the operation of the engine, the desired objectiveis to regulate the flow rate of fuel Q to prevent, within the limits ofpossibility of the combustion chamber, the supercharging pressure fromdescending below a preset threshold; to enable starting up of the turbocompressor unit prior to the engine; and to enable a turned downoperation of the combustion chamber lending itself to rapid resumptionof full power without risk of extinction. Under these conditions, it isdesirable to adopt a relationship for the flow rate of fuel Q to beintroduced into the combustion chamber as a function of the downstreampressure P-ΔP or of the difference between this downstream pressure P-ΔPand the counter pressure P_(c) existing in the regulation chamber 230,such as the relationship illustrated by way of example in the graph ofFIG. 13 in which there is shown as abscissae the supercharging pressureP (relative pressure expressed in bars) and as ordinates the flow rateof fuel Q introduced into the combustion chamber.

In the absence of counter pressure (the pressure in the regulationchamber 230 being equal to atmospheric pressure), the point of operationof the combustion chamber is moved according to this graph. If thispoint of operation is stabilized at point K (idling of the engine), itwill descend toward the point C when the load on the engine increases upto about 20 percent of the maximum load and beyond the point ofoperation is moved between C and D (combustion chamber turned low).

This simple pneumatic displacement of the regulating range can be usedadvantageously to obtain an increase in torque at low speed. It is alsopossible to profit from the displacement of the regulating zone toaccelerate the rise in temperature of the engine (the engine being atidling speed when the combustion chamber is at full power, thermalexchange means being provided between the hot air emerging from thecompressor and the cooling fluid for the engine.

In addition, the introduction of counter pressure which can reach thesupercharging pressure enables the production, by the displacement ofthe regulating zone, of an altimetric compensation. In fact, whenatmospheric pressure diminishes, there can be given to the counterpressure a value equal to the supercharging pressure which has theeffect of opening the second throttle means completely and of supplyingthe combustion chamber at full power (flow rate of primary air and flowrate of fuel reaching their maximum values).

There can hence, due to the invention, be controlled at the same timethe flow rate of fuel injected into the combustion chamber 208 and theflow rate of primary air by acting on the geometry of the secondthrottle means 222 (port 227 and orifice 228a) and on the biasing (andif necessary the adjustment) of the springs 229 and 250. There is thusinsured an air-fuel mixture in the combustion zone 214 in proportionssufficiently close to stoechiometric proportions to obtain, at alloperational speeds of the engine, good stability of the combustionchamber 208.

From the foregoing description it will now be understood that theembodiments described in conjunction with FIGS. 1-8 involve asupercharged engine with a low compression ratio (VR<12) comprising abypass pipe permanently fully open, dimensioned so as to permit thewhole of the flow-rate volume delivered by the compressor to passwithout appreciable pressure drop.

This feature will be evident from the previous description due to thefact that the turbocompressor operates when the engine is stopped, andthat hence the whole of the flow-rate of the compressor passes throughthe bypass pipe. This permanently fully open pipe renders thepermeability of the flow path downstream of the compressor independentof the rotary speed of the engine. This property therefore permits theoperation in stable manner, of the compressor very close to its pumpingor surge line, and therefore in a zone of high yield. Through this factit is possible to reach very high supercharging pressures, taking intoaccount the energy available at the engine exhaust. This ability,associated with the very low compression ratio enables the power of theengine to be considerably increased. However, this widely open pipeprevents any scavenging of the engine.

Accordingly, for engines where scavenging is indispensible (as intwo-stroke engines, or in large bore fourstroke engines due to the heatretention of the valves) the embodiments of FIGS. 9-13 have beenprovided pursuant to the invention to create a pressure-drop-generatingmember arranged in the bypass pipe while, however, preserving theessential property, described above, of the widely open pipe, that is topermit the highly efficient operation of the compressor near its surgeline.

It has been found that this could be achieved by generating a pressuredrop which is independent of the flow-rate passing through the bypasspipe and hence a pressure drop which does not depend directly on therotary speed of the engine. For a given supercharging pressure, theflow-rate of the compressor is given (taking into account this pressuredrop) and the distribution of the flow-rate between the engine and thebypass pipe is organized as a function of the rotary speed of theengine. Therefore, the reduction in this speed does not risk involvingpumping of the compressor.

It also has been found that, taking into account the desirability ofoperating near the surge line, this pressure drop decreases when thesupercharging pressure diminishes and conversely.

In order that the turbine can be driven by the exhaust gases from theengine it is necessary that the relationship between the expansion ratioω of the turbine and the compression ratio ω of the compressor shouldremain substantially constant: ##EQU9##

The regulating valve members described in conjunction with FIGS. 9-13satisfy these two requirements. It also will be understood by thoseskilled in the art from the foregoing description that a scavengingpressure difference may be created by other control devices followingthe principles taught herein, e.g., a movable throttle valve operableunder the control, via suitable balancing and proportioning linkage, offlow sensing and pressure sensing bellows mechanisms to achieve apressure drop in the bypass 106 which, while an increasing function ofonly the pressure existing in the upstream portion of the bypass 106, isnevertheless independent of the air flow rate through bypass 106.

On the other hand, in the embodiment described in conjunction with FIGS.9-13 with adjustable pressure loss, the above functions are ensuredessentially by the construction of the throttle means 108 itself and ina very simple manner, i.e., the throttle means 108 are constituted by amovable valve 108a or 225 cooperating with the fixed seat 108e or 226.This type of means is directly sensitive to the difference in staticpressure generated by the throttle means itself, whilst a flat typebutterfly valve, a turn-clock type valve, etc., are not directlysensitive to such differences in static pressure. The fact alone ofassociating a movable valve with an equilibrating piston working underthe effect of the pressure insures the desired function, viz: theindependence of the flow-rate and the dependence of only the pressureaccording to an increasing function is ensured structurally: ##EQU10##s: cross-section of the equilibrating piston S: cross-section of thevalve

P_(o) : atmospheric pressure or suitable counter pressure

F: stiffness of the return spring

Thus, as in the case of the wide open bypass without any regulatingvalve shown in FIGS. 1-8, the always open but regulated bypass 106insures the independence of the pressure drop generated from theflow-rate in the bypass pipe, and hence independence from the rotaryspeed of the engine thereby enabling stable adaptation of the compressorvery close to its pumping line. This adaptation thus enables very highsupercharging levels to be reached (five and over) with only the energyavailable at the engine exhaust (without a combustion chamber), takinginto account temperatures commonly experienced at the engine exhaust(550° C. to 650° C.).

From the foregoing description, it will now be apparent that theinvention is in no way limited to those of its methods of application,nor to those of its method of production of its various parts, whichhave been more especially indicated; it encompasses, on the contrary,all variations. However, by way of further illustration and not by wayof limitation, an actual working embodiment incorporating the presentinvention, along with its operational and performance characteristicsand a comparison to the prior art, will now be described and illustratedin conjunction with FIGS. 14-17.

FIG. 14 is a schematic diagram of the engine and supercharging systemshown in FIGS. 15 and 16 and incorporates elements of the systems shownand described in conjunction with FIGS. 9-14 as well as certainadditional features which are the subject of my U.S. Pat. Nos.3,849,988, 3,894,392 and 3,949,555. In practicing the method ofsupercharging of the invention as described previously, a conventionalcommercially available diesel engine 1, specifically the Poyaud Model520-6L S3 engine manufactured by Societe Surgerienne de ConstructionMecaniques of Surgeres, France, was modified to contain or be equippedwith the following elements: a starting unit 300 corresponding to unit 6described previous, a compressor 302 and turbine 304 together forming aturbo-compressor unit similar to unit 2-4-5 described previously, acombustion chamber 306 corresponding to chamber 3 or 208, a fuelinjection unit 308 similar to unit 7 or units 210 or 210a and associatedsystems of FIGS. 10 and 11 respectively, an exhaust manifold 310corresponding to manifold 1b, a bypass regulator 312 of the type shownin FIGS. 9, 10 or 11, a bypass pipe 314 arranged as shown in FIGS. 2 or9, an adjustable preheating valve 316, an air cooler 318 arranged asdisclosed in my copending application Ser. No. 263,759, and a suitablecontrol unit 320 for correlating the functions of starter 300, regulator312 and injection unit 308 in accordance with the previously describedmethod of the invention. The manner in which the aforementioned elementsare structurally incorporated with the aforementioned Poyaud Model520-6L S3 engine is shown in FIGS. 15 and 16 wherein like referencenumerals denote the corresponding elements.

The starting unit 300 insures the starting and lubrication of theturbo-compressor 302-304 and the supply of fuel to the combustionchamber 306. This unit ceases to operate once the engine has startedunder its own power whereupon the functions of lubrication and fuelsupply as required to the combustion chamber 306 are assumed by theengine oiling system and an engine driven fuel pump. The combustionchamber 306 contains the injection unit 308 and is mounted in theexhaust manifold 310 which in turn is suitably adapted to therequirements of a given engine as will be understood by those skilled inthe art from the foregoing description. The combustion chamber 306constitutes the thermodynamic heart of the hybrid machine formed by theinvention and organizes the mixture of the exhaust gases leaving theengine 1 and the bypassed air from conduit 314. The regulator 312 isflanged to the upstream end of the exhaust manifold 310 and performs thefollowing functions: (a) control of the pressure loss or pressure dropin bypass 314 and hence the scavenging ratio; (b) proportioning of theair flow to the combustion chamber 306; (c) proportioning of the fuelflow to the combustion chamber 306; (d) regulation of the turbine 304 toa pressure level displayed on a control board associated with control320 for economy or extra torque or speed modes of operation; and (e)altimetric correction. The preheating valve 316 bypasses the air cooler318 during heating of the cooling water preparatory to start up, and thecontrol unit 320 centralizes all of the functional commands withsuitable safety factors providing a system of high reliability.

The Poyaud engine equipped as described above is operated with no moredifficulty than a conventional diesel engine, the preliminary startingup of the compressor 302 being completely automatic. The engine may thenbe started independently of the ambient temperature conditions and thetime required for heating up the cooling water is much reduced, thecooler 318 for the supercharging air initially operating as a waterheater. Moreover, the operator has available the control 320 whichenables him to obtain extra torque instantaneously from the enginewhenever such is needed. This enables considerable reduction in thenumber of gear box ratios and in practice may be operated somewhat likethe conventional overdrive which it replaces. The permanent existence ofan excess of air being fed to the engine and chamber 306 avoids anyfouling of the engine combustion chambers as well as of the distributorof the turbine 304.

Referring more particularly to the engine shown in FIG. 15, whichillustrates the first application of the invention to a commerciallyavailable engine, and to the test results shown in FIG. 17 which is aplot of engine rpm against horsepower (lower graph) and against specificfuel consumption of grams per horsepower hour (upper graph), it will beseen how the horsepower of this engine was increased when it wasmodified according to the present invention, as demonstrated bycomparative tests carried out on a conventional production line dieselengine (A) unsupercharged, (B) supercharged conventionally and (C)modified according to the present invention.

The Poyaud 520-6L S3 engine develops 180 hp at 2500 rpm in itsnonsupercharged version (A), and 300 hp at 2500 rpm in itsconventionally supercharged version (B) with cooling of thesupercharging air.

The adaptation of the engine with the supercharged system of theinvention (C) enables this power to be brought to 600 hp in continuousservice.

This adaptation is effected on a production line engine by means of thefollowing modifications:

1. Pistons--The combustion chamber of the pistons of the production lineengine is remachined so as to reduce its volumetric ratio from 15 to8.55.

2. Injection pump--The diameters of the pistons of the injection pumpwere brought from 90 to 130 mm so as to be able to double the amount offuel injected.

3. Replacement of the conventional turbo-blower having a 2.4 pressureratio by a turbo-blower 302-304 with a pressure ratio of 4.8 andincluding self-starting means.

4. Replacement of the production line intake manifold, which has one aircooler, with a manifold comprising two air coolers 318 with the samecharacteristics.

5. Replacement of the production line exhaust manifold by a specialexhaust manifold 310 including the combustion chamber 306.

6. A bypass pipe 134 (3.5 inches diameter) connects the intake manifolddirectly to the exhaust manifold.

7. A pipe normally closed by a butterfly valve 316 and enabling thebypass of the coolers for the supercharging air and brought into actionduring the phase of starting up and of increasing temperature of theengine.

8. An auxiliary unit 300 enabling self-sustaining operation (especiallywhen the engine has stopped) of the supercharging turbocompressor. Thisassembly includes especially: (a) an electropump for oil enablinglubrication of the turbo-compressor before the starting of the engine;(b) an electropump for fuel supplying the combustion chamber; (c)devices enabling the starting of the turbocompressor and the ignition ofthe combustion chamber.

9. A fuel regulator 308 for the combustion chamber. This regulator isservocoupled to the supercharging pressure. Its principal function is toprevent the supercharging pressure from falling below a threshold levelwhich is a given value based upon the value of the volumetric ratio (1.8bar for a volumetric ratio 8.55) as well as the other parametersinvolved in establishing self-ignition conditions in the engine.

This adaptation is produced on a production line engine withoutmodification of the main parts of the engine (lower engine, crankshaft,linkage, engine block, cylinder heads, and distributor, oil pumps,auxiliary members and cooling systems for water and oil).

The 600 hp power (curve C) is obtained with the same maximum combustionpressure, and exhaust temperature as in the conventionally superchargedengine (B) operating at 300 hp (135 bar-600° C.).

The specific consumption at maximum power is practicallyunchanged--(compare the Table of performances and curves A, B and C ofthe upper graph of FIG. 17). On the other hand, the consumption at lowspeed is slightly greater than that of the production line engine. Whenoperating engine C at over about 20 percent of its maximum power, thecombustion chamber 306 operates at pilot level and its fuel consumptionis negligible. Below about 20 percent of the maximum power, the fuelflow increases and the consumption of the combustion chamber is added tothat of the engine. This data is shown in the following table and alsoon the graph of FIG. 17 which shows power and specific fuel consumptionas a function of engine speed (at maximum torque).

    ______________________________________                                        Technical Characteristics                                                               Production line engine                                                                          Invention                                                   Nonsupercharged                                                                          Supercharged                                                                             Version                                                   (A)        (B)       (C)                                          ______________________________________                                        Cylinder                                                                      stroke volume                                                                             10.47 l      10.47      10.47                                     Weight      970 kg       1040       1075                                      Volumetric                                                                    ratio       15           15         8.55                                      Supercharging                                                                 pressure    1 bar        2.4        4.8                                       Maximum                                                                       pressure    --           135 bar    135                                       Mean effective                                                                pressure at                                                                   2500 rpm    6 bar        10         21                                        Power at con-                                                                 tinuous speed                                                                 of 2500 rpm 180 hp (metric)                                                                            300        600                                       Specific con-                                                                 sumption of                                                                   fuel at maximum                                                               power       (182g/hp/h   175        175                                       Specific flow                                                                 rate of air                                                                   at maximum                                                                    power       4.5kg/hp/h   5.2        6.2                                       Temperature                                                                   at the outlet                                                                 of the cylinder                                                               head at maximum                                                               power                    580° C.                                                                           600° C.                            ______________________________________                                    

With respect to the anti-pollution aspects of the present invention, itshould first be noted that the principal polutants emitted by aninternal combustion engine are:

NO, NO_(x) : oxides of nitrogen

CO: carbon monoxide

C_(x) H_(y) : unburnt hydrocarbons

In a lean mixture, the unburnt hydrocarbons are very low, evennegligible.

Numerous measurements (see Wimmer, D.B., and McReynolds, L.A., "NitrogenOxides and Engine Combustion," Paper 380E, SAE Summer Meeting, St.Louis, MI, June 1961; Alperstein and Bradow, "Exhaust Emissions Relatedto Engine Combustion Reactions," SAE Transactions, Vol. 75 Paper 660781,1967; and Huls, T.A., and Nickol, H.A., "Influence of Engine Variableson Exhaust Oxides of Nitrogen Concentrations from a Multi-CylinderEngine," SAE Paper 670482, May 1967) indicate that the concentrations ofthe pollutant species (principally NO and CO) correspond in practice toequilibrium concentrations calculated at the point of maximumtemperature and not to those calculated at the temperature of theexhaust. This is explained by the unbalanced behaviour of thesepollutant species during combustion and expansion. During combustion, NOand CO are formed very rapidly and reach equilibrium levels at thetemperature at which combustion occurs. During expansion, the pollutantspecies are decomposed very slowly. Thus the formation of nitric oxideand of carbon monomide is fixed in practice by the maximum temperatureof the cycle.

The rates of formation K_(f) and of decomposition K_(d) of these speciesare in the ratio of the equilibrium constant K_(c) of the reactionconsidered as follows: ##EQU11##

The interpretation of the phenomenon of "freezing" at the maximumtemperature of the cycle has been given by numerous researchers (seeNewhall, H. K., and Starkman, E. S., "Direct Spectroscopic Determinationof Nitric Oxide in Reciprocating Engine Cylinders," Paper 670122, SAEAutomative Meeting, Detroit, Mich., Jan. 1967; Newhall, H. K., "Kineticsof Engine-Generated Nitrogen Oxides and Carbon Monoxide," TwelfthSymposium [International] on Combustion, The Combustion Institute, 1969;Eyzat, P., and Guibet, J. C., "A New Look at Nitrogen Oxides Formationin Internal Combustion Engines," SAE Paper 680124, Jan. 1968; Lavoie, G.A., Heywood, J. B., and Keck, J. C., "Experimental and Theoretical Studyof Nitric Oxide Formation in Internal Combustion Engines," CombustionScience and Technology, Vol. 1, Feb. 1970; and Spadaccini, L. J., andChinitz W., "An Investigation of Non-Equilibrium Effects in an InternalCombustion Engine," Trans. of ASME, April 1972, pp. 98-107).

Cycles at low compression ratio and high supercharging rate are coldcycles. Thus comparing the maximum temperature of the cycle calculatedfor the two following different cycles, for example:

    ______________________________________                                        Compression ratio  13.1       7                                               Supercharging pressure                                                                           2.75 bar   7 bar                                           Intake temperature 65° C.                                                                            65° C.                                   Temperature at exhaust outlet                                                                    1000° C.                                                                          1000° C.                                 Maximum pressure of cycle                                                                        140 bar    140 bar                                         Maximum temperature of cycle                                                                     2400° K.                                                                          1780° K.                                 Effective average pressure                                                                       16 bar     32 bar                                          ______________________________________                                    

Hence reduction of the compression ratio and relative increase in thesupercharging rate reduce the maximum temperature of a cycle by morethan 600° C. The formation of NO_(x) and CO would thereby be expected tobe considerably reduced.

Actual tests were conducted on an assembly according to the presentinvention using an engine modified pursuant to the invention asdescribed above. The results were as follows:

    ______________________________________                                        Measurement of Exhaust Emission of Engine (C)                                                NO.sub.2                                                                             CO       He                                                            (g/hph)                                                                              (g/hph)  HC                                             ______________________________________                                        (1°) At 2500 revs/min                                                   300 hp          7.2      0.8      <1                                          425 hp          6.3      1.8      <1                                          600 hp          6.6      4.3      <1                                         (2°) At idling speed of engine (combustion chamber at                  maximum rate)                                                                 ______________________________________                                    

As percentage of discharge of dry exhaust

CO: 0.16%

NO₂ : not measurable

The 520 engine (C) is a direct injection engine (combustion chamberlocated in the piston).

It will be seen that the level of pollutants from engine (C) is lessthan half that shown for a direct injection engine, namely about 14ghph, in FIG. 3 of the study by C. J. Walder published in Ingenieurs del'Automobile 7/8-72, p. 407, "Moteur et Carburants".

The improved results obtained from the foregoing example of a Poyaudengine supercharged in accordance with the principles of the presentinvention are believed to be primarily due to several features whichcooperate in the combination of the method and system of the inventionto provide the substantial increase in power-to-weight ratio without asignificant increase in specific fuel consumption as well as the manyother advantages indicated previously as well as hereinafter. Althoughall of the theoretical bases for the improved results may not as yet befully understood, it is believed that a discussion of some of thetheoretical aspects of the invention may be helpful in providing abetter understanding of why such results are achieved and of how theinvention may be best applied to other and more diverse engine systems.

As indicated previously, as a first approximation, it should berecognized that the power of a diesel engine is proportional to the flowof air passing through it. The intake volume being limited by the linearspeed of the pistons, it is necessary to increase the density of theintake air; i.e., to increase the pressure and decrease the temperatureof the air up to the limits stipulated by chemical kinetics and theengine structure. These factors have led engine makers to equip engineswith compressors and air coolers for the intake air. In one type ofprior art supercharging system the compressors are connected to theengine crankshaft. However, this solution cannot be used in the case ofa highly supercharged engine where the compression power may reach andexceed the power of the supercharged engine. Consequently, the presentinvention is concerned primarily with improvements in another type ofprior art system, namely, turbo-supercharging systems whereinturbocompressors are powered by exhaust gases.

Turbo-supercharging first appeared before the Second World War and wasadopted on a large scale only after 1950. The first turboblowers hadpressure ratios too low to make an appreciable change in thethermodynamic cycle of engines, the structure of the engine remainingunchanged. For a considerable time, supercharged engines were naturallyaspirating engines equipped with a "pressure booster," considered to bean optional accessory. The gradual increase in pressure ratio was thecause of difficulties which brought considerable discredit onsupercharging, which is still synonymous with fragility in certaincircles. Subsequently, engines were designed to withstand supercharging,and until recent years the technologies of engines and of blowers haveevolved in parallel. More recently, the progress made in centrifugal andaxial compressors have provided increases in air pressures whichconventional engines have been incapable of utilizing. Therefore, toachieve further progress, the hybrid assembly formed by the engine, theturboblower and the cooling system must be re-evaluated. Accordingly,consideration will first be given to the problems raised bysupercharging of engines and their consequences. With respect to theeffect of supercharging on mechanical stresses, it first should be notedthat a diesel engine is dimensioned for the maximum stresses it willwithstand in the region of the combustion dead center; i.e., during avery small fraction of its cycle. Let P_(o) (α) be the development ofpressure as a function of crank angle of the nonsupercharged cylinder.Let π be the supercharging ratio=ratio of intake pressure of thesupercharged cylinder to that of the nonsupercharged cylinder. Thecompression phase of the supercharged engine is in accordance with theequation:

    P.sub.l (α)=P.sub.o (α)

The supercharging rate of a given engine is obviously limited to:##EQU12## π max is about 4 for large bores. If we wish to respect P maxand increase π max, P_(o) comp., i.e., the compression ratio of theengine, must be reduced. Thus, an initial result can be established;i.e., the supercharging rate can be increased without changingmechanical stresses by decreasing the volumetric ratio ε, as will beseen by the comparative curves shown in FIG. 18.

With respect to the effect of supercharging on the thermal load, itshould first be noted that the thermal load is measured by thetemperature of the piston wall, at the top of the lining and at thebottom of the cylinder-head. It can be shown that these temperaturesincrease linearly with the thermal flow passing through the metal at agiven point. Mathematical expression of this thermal flow is verycomplicated, and therefore it is believed for present purposessufficient to merely reason by comparison between more or lesssupercharged cylinders.

Let T_(o) (α) be the development of temperature as a function of thecrank-angle of the nonsupercharged cylinder and θ the ratio of absoluteintake temperatures of supercharged and nonsupercharged cylinders. Thecompression phase of the supercharged cylinder is given by:

    T.sub.l (α)=θT.sub.o (α)

As in the case of pressure, it is observed that the temperature at theend of compression is multiplied by θ. This temperature is importantbecause it fixes initiation level of combustion. It is difficult toreduce θ below 1.25 without a considerable increase in the volume ofcoolers (the cold source being the atmosphere). T_(o) comp. musttherefore be reduced and this is a second reason for reducing thevolumetric ratio ε. To provide a practical illustration, thetemperatures at the top of the lining of the same cylinder may betheoretically compared in the:

(a) conventional supercharging π=3, volumetric ratio 12

(b) invention supercharging π=6, volumetric ratio 7 with excess airgiving the same temperature at the end of expansion as at (a).

    ______________________________________                                                       Conventional                                                                             Invention                                           ______________________________________                                        Volumetric ratio  12           7                                              Intake pressure   3            6                                              Intake temperature                                                                              65° C.                                                                              80° C.                                  Compression pressure                                                                            90 bar       85 bar                                         Compression temperature                                                                         577° C.                                                                             452° C.                                 Combustion pressure                                                                             135 bar      135 bar                                        Combustion temperature                                                                          1627° C.                                                                            1427° C.                                Pressure at the end of                                                                          10.5 bar     20 bar                                         expansion                                                                     Temperature at the end                                                                          900° C.                                                                             900° C.                                 of expansion                                                                  ______________________________________                                    

To compare the temperature T_(p) of the wall at the top of the lining ofthese two cylinders, the density of the heat flow density Φ passingthrough the lining at this point is calculated. We know that during thetime dα, the transfer per unit area E at the point concerned, is givenby the equation:

    dE=Φ(α)dα=h(α)[T(α)-T.sub.p ]·dα

and ##EQU13## in which Φ is the average heat flow density and h (α) isthe transfer coefficient which depends on the temperature and thepressure.

Referring to the set of graphs in the accompanying drawings, FIG. 18gives the comparative theoretical developments of pressure in theconventional cycle and in the cycle of the invention, FIG. 19 gives thecorresponding theoretical temperature developments, FIG. 20 gives thecorresponding theoretical developments of the transfer coefficient, andFIG. 21 gives a comparison of the theoretical heat flow density at thepoint concerned for a wall temperature of 260° C. It will be seen thatthe top of a cylinder in an engine of the invention developing a meaneffective pressure (M.E.P.) ranging about 30 bars is cooler than that ofthe same conventional cylinder at an average pressure of 20 bars. Thisis chiefly due to a decrease in combustion temperatures which influenceto the fourth power the energy radiated toward the walls.

With respect to the thermal balance of the engine cycle of the inventionand its effects on the cooling system, reference may be made to thefollowing tables which give a comparison between the theoreticaltemperatures (Table 1) and theoretical heat flows (Table 2) of the sameengine in conventional supercharging and in supercharging pursuant tothe invention:

                  TABLE 1                                                         ______________________________________                                        Temperatures   Conventional                                                                              Invention                                          ______________________________________                                        Supercharging pressure                                                                       3 bar       6 bar                                              Air consumption                                                                              5.2 Kg/H.P./H                                                                             6.9 Kg/H.P./Hour                                   Mean effective pressure                                                                      17 bar      30 bar                                             Intake temperature                                                                           15° C.                                                                             15° C.                                      Compressor output                                                                            165° C.                                                                            270° C.                                     temperature                                                                   Engine intake  60° C.                                                                             80° C.                                      temperature                                                                   Turbine intake 580° C.                                                                            600° C.                                     temperature                                                                   Turbine output 450° C.                                                                            350° C.                                     temperature                                                                   ______________________________________                                    

                  TABLE 2                                                         ______________________________________                                        Thermal Balance (in % of fuel enthalpy)                                       ______________________________________                                        Brake power          36.5%    36.5%                                           Power removed by cooling air                                                                       7.5%     15.5%                                           Power losses in water and oil                                                                      16.5%    9.5%                                            Exhaust losses       33%      29.5%                                           Other losses         6.5%     9%                                              ______________________________________                                    

It will be seen from Table 2 that 62% (15.5/15.5+9.5) of the heat to beextracted is in the supercharging air, with the invention, and only 31%(7.5/7.5+16.5) with conventional supercharging. These calories are athigh temperature, so they are easily evacuated if a direct exchange ismade with the cold source (atmosphere or external water). The averagetheoretical temperature differences are given in the following Table 3:

                                      TABLE 3                                     __________________________________________________________________________    Fluid to be Cooled                                                                             Conventional                                                                            Invention                                          __________________________________________________________________________    Cylinder block cooling water                                                  and lubricating oil -                                                         Input/output temperature of  coolant                                                            ##STR1##                                                                                ##STR2##                                          Mean temperature difference                                                   with cooling air at 40° C.                                                                 47° C.                                                                           47° C.                                   Power to be extracted as % of                                                 engine power      45%       26%                                               Size of cooler   91        55                                                 Supercharging air -                                                           Intake/output cooler temperatures                                                               ##STR3##                                                                                ##STR4##                                          Mean temperature difference with                                              cooler fluid (air at 40° C.                                                                70° C.                                                                           135° C.                                  Power to be extracted as % of                                                 engine power      25%       43%                                               Size of cooler   30        32                                                 Overall size of coolers                                                                        121       87                                                 __________________________________________________________________________

To summarize the above considerations, it will be seen that thereduction of the compression ratio has three favorable effects on thethermal load: (1) at equal temperatures in the intake, it decreases theheating of the air during compression and, therefore, the temperature atthe beginning of combustion; (2) for given values of maximum pressure,brake mean effective pressure and intake temperature, it enables theexcess of air to be chosen to reduce the combustion temperature and theheat transfer by radiation; and (3) at equal supercharging pressures, itreduces all pressure levels and, therefore, the value of the gas to wallheat transfer coefficient.

Another factor to consider in supercharging is the choice of intaketemperature. Chemical kinetics imposes end of compression conditionssuch that the ignition delay must be sufficiently short under alloperational conditions and in particular during operation at no load.This delay depends on the pressure and temperature of compression whichin turn depend on the conditions of intake and on the volumetric ratio.Below a certain volumetric ratio, which varies with the bore (12 for thelargest, 17 for the smallest), the intake pressure and temperature musttherefore be artificially maintained above ambient conditions in orderto ensure a good ignition of the fuel in the engine. It is possible toact upon the temperature alone, or on the pressure alone or on bothsimultaneously.

The effect of temperature alone is relied upon in the conventionalheating of the intake manifolds in order to facilitate starting. Forvery low volumetric ratios, the compression pressure may drop below 10bars. The necessary intake temperature for correct ignition may thenexceed 300° C. The density of intake air is then halved and this has twodetrimental effects: (1) an increase of intake temperature facilitatingignition is accompanied by a decrease in the density of oxygen which isunfavorable to ignition; and (2) the average pressure may becomeinadequate to accelerate the engine.

Since every compression is accompanied by heating, the case of theeffect of pressure alone has to be considered only if the air coolersare very cold, which is the case on starting. Starting is then onlypossible with an intake pressure requiring an oversized startingmechanism.

However, in accordance with the present invention, there is asimultaneous effect of temperature and pressure in obtaining the properintake temperature. This is very simply achieved by momentary bypassingof the air coolers during starting and heating periods of the engine, asset forth in more detail in my U.S. Pat. No. 3,894,392.

In considering the performances possible with the supercharging systemof the invention, the performances have been calculated for an enginedesigned to have a combustion pressure of 135 bars and a temperatureinside the cylinder of 900° C. at the time of exhaust opening. A brakemean effective pressure greater than 30 bars may be expected in the nearfuture. FIG. 22 shows the theoretical indicated mean effective pressure(I.M.E.P.) as a function of the supercharging pressure and of theangular duration of combustion. This diagram reveals the considerableinfluence of rotary speed on the average pressure of diesel engines.Thus, high speed engines are very limited by chemical kinetics. However,it is to be noted that the excess of air in the cycle of the inventionis used to accelerate the end of combustion, thereby tending to offset adrop in the indicated mean effective pressure with increasing enginespeed.

With respect to the effects of supercharging pursuant to the presentinvention on the mechanical operation of the engine, it will beunderstood that the combustion diagram is fairly comparable with that ofa spark ignition engine and hence gives the engine smooth operation anda low ratio between maximum and average pressure. Moreover, the specificlosses by friction are reduced in the same ratio as the mean effectivepressure. As a result, the mechanical efficiency is increased by about10 points.

The effects of supercharging pursuant to the present invention onspecific fuel consumption may be considered surprising because it isgenerally assumed that a reduction in the compression volumetric ratiois necessarily accompanied by an increase in fuel consumption. This isnot true, the proof being that a supercharged engine consumes less thanin the aspirating version, in spite of its lower volumetric ratio.Without going into a strict study of thermal efficiency, it should bepointed out that factors favorable to supercharging in accordance withthe present invention balance the unfavorable effect of the reduction inthe engine compression ratio. For example, the specific losses in theengine cooling water and oil are approximately halved, the mechanicalefficiency is increased by about 10 points, and the specific losses byradiation and convection are approximately halved. Although at thepresent time there is a lack of experimental results on very lowvolumetric ratios, it is probable that supercharging pursuant to thepresent invention does not appreciably change the specific fuelconsumption even at such very low ratios.

As briefly indicated previously, optimum results from the superchargingsystem of the invention are obtainable only if the compressor operatingpoint moves along a line very close to the surge zone. Such a mode ofoperation of the compressor is impossible if the engine were to beconnected in series between the compressor and the turbine. It is known,in fact, that the permeability of the engine varies to a very greatextent according to running speed, load, overlap time, etc. In addition,the transient operations of the engine associated with the inertia ofthe blower rotor frequently entail surging of the compressor which isnot desirable for high pressure engines.

The system of the invention avoids this difficulty by controlling thepressure drop between the compressor and the turbine by means of anengine bypass duct such that the permeability of the compressor-turbineair flow circuit is rendered independent of the engine operation. Thisconcept also assumes the use of a supercharging system of the constantpressure type, which is in any case necessitated by consideration of theefficiency of expansion. The simplest version is a widely open tubewithout any structure for varying its flow characteristics, i.e., asimple pipe of fixed geometry, as shown and described in conjunctionwith FIGS. 1-8 which is large enough in cross section to produce noappreciable pressure loss. However, this attractive solution does notpermit scavenging of the engine, and therefore for scavengingapplications the modifications shown and described in FIGS. 9-13 arepreferably employed.

From the foregoing description, it will be understood that the choice ofvery low volumetric ratios requires that intake pressure and temperaturenever decrease below certain values which are higher than ambientvalues. However, in the system of the invention, these conditions arefulfilled between the compressor 2 and the combustion chamber 3 of thegas generator of the turbine 4 whose idling is correctly adjusted. In asense, then, pursuant to the present invention supercharging consists inmaking the engine "breathe" in this artificial atmosphere in the sameway as a deep-sea diver in his diving suit.

The supercharging system of the invention assumes that for maximumbenefits a very high intake pressure, 6 to 7 bars, will be in common usein the near future. At such pressure ratios, the energy available in theexhaust gases at 600° C. will not be sufficient to drive the compressorunless the isentropic efficiencies of compression and expansion are highand the pressure loss between the compressor and the turbine ismoderate. It is also known that a transonic or supersonic compressor hasa good yield only in the immediate vicinity of its surge line as shownby the characteristic field of FIG. 23.

Thus, such elevated boost pressures achieved by the present inventionare not possible unless the compressor operates very close to maximumefficiency. It is not simply a matter of finding a pump which will givehigh enough pressures. Rather, it goes to the basic operation of thecompressor. To achieve the necessary pressures, the compressor mustoperate at about 75% of efficiency, and what this means is that the pumpmust be permitted to work very close to its "pumping" or surge line. Inorder to understand why it is not practicable to use turbocompressorswith very high compression ratio (of the order of 5 to 7 or more)without operating very close to the surge line, it first should be notedthat, as indicated above, if the isentropic yield of the compressor is0.8 (close to the pumping line) the autonomous temperature at the inputof the turbine, enabling the compressor to be driven, is of the order of600° C. This relatively low input temperature condition of a compressoroperating at high efficiency can be obtained with the diesel enginealone, without the outside addition of energy to the combustion chamber(the rotary speed of the turbine remains limited: N=80,000 revs/min forexample). However, if a turbocompressor cannot be operated close to thesurge line, such as for reasons of stability of the turbocompressor,i.e., pumping in the case of sudden reduction in the permeability ofwhatever passageway or passageways are provided between the compressorand turbine, the yield will be bad, for example n=0.6. In order to drivethe compressor at this lower yield factor, an input temperature at theturbine of about 900° C. is required in order to provide the energynecessary at the turbine. This thermal level cannot be obtained with thediesel gases alone because the cylinder head, pistons and especially theexhaust valves could not withstand this temperature level. Thus, thenormal maximum temperature of the exhaust gas is between 550° C. and650° C. It would therefore be necessary to continuously heat the dieselexhaust gases with a combustion chamber in order to raise thetemperature of the gases from 600° C. to 900° C. or higher. Increasingthe heating of the exhaust gases by 300° C. would increase by 20% theconsumption of fuel of the installation. This, of course, would be veryexpensive and would have to be supplied to the nominal power of theengine, that is to say, throughout its use. In addition, the speed ofthe turbine would be very high (of the order of 100,000 revs/min in theabove example), and the higher turbine inlet temperature would requiremore expensive materials.

Thus, to operate at a very high pressure level with energy derived onlyfrom the diesel gases, taking into account the permissible heat level atthe engine exhaust (about 600° C.), it is necessary to operate with avery good yield (of the order to 75%) and thus very close to the surgeline. FIG. 23 illustrates diagramatically the operating characteristicsof a compressor suitable for supercharging. The pump has a surge lineand the closer the pump can operate to that surge line the greater itsoperating efficiency. However, once the pressure conditions exceed (moveto the left of) the surge line, there is a back flow which totallydisrupts the smooth flow of air through the compressor.

It must now also be appreciated that a diesel engine connected in seriesto a compressor acts like a throttling restriction when during runningthe speed of the engine suddenly moves from high speed to low speedconditions. The result in such prior art systems is surging at thecompressor which can easily damage the compressor and the engine. Inrecognition of this problem, the prior art has operated far to the rightof the surge line, as indicated by curves A and B. Curve A representingacceleration and curve B representing deceleration. In contrast thereto,however, the present invention operates close to and parallel to thesurge line without danger of surging, as shown at curve C, due to theprovision of the constantly open gas flow bypass between the compressorand the turbine. Since, as described in conjunction with FIGS. 1-8, thisbypass must clearly provide a passage for the output of theturbocompressor on starting the latter, without appreciable pressurelosses, it must obviously be a wide bypass; i.e., large enough in crosssection to avoid acting in any appreciable manner as a fixed throttlingorifice. In this way, variations in engine speed do not substantiallyaffect the air flow-rate through the compressor. Rather, this bypass,which is always open, instantly accepts a greater portion of the airflow leaving the compressor when the engine speed is reduced, and afortiori when it is suddenly drastically reduced, so that the turbineinput or compressor output can remain essentially unobstructed, therebypermitting smooth high-yield operation of the compressor near its surgeline.

Although the prior art shows various bypasses in supercharging systems,none of these show a bypass having the permeability property of theabove-described bypass passageway of the present invention, perhapsbecause it was mistakenly thought that a direct constantly open bypasswould render the aerodynamic flow so unstable that the diesel enginewould reaspirate its exhaust gases through the bypass. However, it hasbeen found that such a condition does not occur. To the contrary, it hasbeen found that an always-open bypass in which the pressure drop isindependent of air flow in the bypass can alone ensure the aerodynamicadaptation of two so incompatible machines as a volumetric machine(diesel engine) and a turbo machine (turbo-compressor) when thesediverse machines are properly matched pursuant to the invention. It willbe understood that the invention, whether applied to scavenged ornonscavenged engines, assumes that the turbine-compressor superchargerhas been properly selected or matched to the engine to provide for anycondition of engine operation at least some air flowing from thecompressor to the turbine via the bypass branch even at maximum air flowcondition through the engine combustion chamber system.

The invention will be better understood from the following descriptionof the operation with reference to the relationship between boostpressure versus percentage of maximum power. Owing to the fact that thelow compression ratio engine will not develop sufficient pressure withatmospheric air for self-ignition at starting and low power conditions,the turbocharger is first started (by, for example, the separatestarting means 6 shown schematically in FIG. 1). At this time the airfrom the compressor flows only through the bypass (which is, of course,very wide) since the engine has not yet started. Fuel is delivered tothe combustion chamber for operating the turbine so that there isdeveloped at the compressor outlet a pressure of 2 atm which in thepresent instance is the boost pressure sufficient for self-ignition atstarting and low speed conditions. At this point, the engine is startedand this 2 atm pressure is delivered to the intake manifold of theengine whereby self-ignition now commences. In the example, the boostpressure of 2 atm represents the minimum or "threshold" value belowwhich the boost pressure should not be permitted to fall lest thepressure at the engine be insufficient for self-ignition at starting andlow power. This threshold value will depend on the structuraldimensions, i.e., the compression ratio and bore, of the engine and willbe fixed for a given engine. Moreover, this threshold value of pressurerepresents a certain speed of the turbine, and hence maintaining aminimum threshold value of boost pressure is directly translated intopreventing the turbine speed from falling below a known minimum numberof revolutions per minute.

In the preferred embodiment, maintenance of the threshold pressure valueis thus accomplished by delivering sufficient fuel to the combustionchamber 3 to heat the gases entering the turbine a sufficient amount tomaintain the threshold pressure of the gases and hence the correspondingthreshold speed of the turbine. This threshold value is maintainedessentially constant from zero to about 20% of full power. Additionalfuel could continue to be delivered to the combustion chamber 3 whilethe boost pressure rises continuously from 2 atm at zero power to 4.8atm at 100% power, but this would be a waste. As noted, the fueldelivered to the combustion chamber 3 provides sufficient heat in thebypass to maintain the gases entering the turbine at a sufficienttemperature to give a compressor outlet pressure of 2 atm. However, asthe engine moves from 0 to 20% of full power, hot gases start to leavethe diesel engine and enter the turbine. Thus, the amount of fuel to thecombustion chamber 3 required to maintain the 2 atm boost pressure dropsoff from an initial level at zero power (before the engine has started)to zero at 20% of full power. Above about 20% of full power, if theefficieny of the turbocompressor assembly is high enough, the exhaustgases from the engine are sufficient to maintain and increase theturbine speed and hence also the boost pressure without additional fuelin the combustion chamber 3, resulting in rising boost pressure withincreasing power. It is, of course, apparent from the previousdescription that variations in the fuel supply need not be continuouslydependent on the turbine speed. The only essential function of theregulating device 7 used in the present invention is to assure that thecompressor outlet does not fall below a predetermined minimum pressurewhich in turn means insuring that the turbine speed does not fall belowa minimum predetermined speed which minimum speed is fixed for a givenengine, all other parameters affecting self-ignition being keptconstant. Thus, taking such parameters into account, this means that byway of example, fuel must be delivered to the combustion chamber 3 insufficient amount to assure that the turbine does not fall below 50,000r.p.m. Therefore, this amount could be a constant amount but withunnecessary waste of fuel which could be avoided by a regulating device.

Regulating devices for relating compressor pressure or turbine speed toa combustion chamber fuel supply are conventional and hence will be wellknown to one skilled in the art. For example, the U.S. Pat. Nos. toPrince 2,379,455, Nettel 2,654,991 and Zuhn 3,096,615 describeregulating means which respond to the pressure delivered by thecompressor. The Nettel U.S. Pat. No. 2,620,621 and the Dumont U.S. Pat.No. 3,163,984 disclose regulating means responsive to the temperature atthe input of the turbine. However, an operative "link" between theturbine-compressor supercharger and the combustion chamber as disclosedin conjunction with FIGS. 10-13 herein are preferred, at least insupercharging systems of the invention requiring scavenging.

The operating sequences of the hybrid assembly of the invention can alsobe graphically summarized as shown in FIG. 24. First, as explainedpreviously, the gas turbine is started up and made self-sustaining withthe combustion chamber 3. Then the diesel engine 1 is started byconventional means (electric or pneumatic starter) and while idlingprovides no significant power to the turbine. However, once the dieselengine is placed under load, there can be three general phases.Referring to FIG. 24, there is a constant supercharging pressure phase.This phase corresponds approximately to the range of power between 0 and20% of the maximum power at which the engine takes over from thecombustion chamber. Secondly, there is a variable supercharging pressurephase. In this zone, between 20 and 100% of its power, the engine worksas a gas turbine combustion chamber whose pressure loss is controlled.Thirdly, for traction applications where high torque at low runningspeed is desirable, a maximum supercharging pressure phase may beobtained by an overriding order to the fuel regulator.

From the foregoing description it will now be apparent that oneimportant advantage of the system of the invention is that it can beapplied to almost all existing engines. The transformation of aconventional engine to the method of the invention essentially affectsits environment and its equipment. The exhaust manifold is thethermodynamic heart of the hybrid machine. It contains the combustionchamber and its control parts. Thus, at the cost of minor modifications,the invention multiplies the power of nonsupercharged engines by afactor between 3.5 and 4 in the present state of technology withoutreducing the life of the engine. Moreover, anticipated progress fromcurrent work to extend the limits of conventional supercharging will bedirectly usable by supercharging systems of the invention and with anincrease in the relative gain.

Since the supercharging method and system is dependent on the existenceof high pressure turboblowers for maximum benefits, it is anticipatedthat the invention will be the main industrial application of recentdevelopments in supersonic centrifugal and axial compressors. In thisrespect, turbocompressors must now be classified by the type of turbinewith which they are equipped. The radial turbine, capable of a highexpansion rate, is well adapted to the supercharging system of theinvention. Unfortunately, its gas flow is limited and hence it is usableonly for small power units. Today there are a few machines in thiscategory whose pressure ratio exceeds 5. One of such radial turbines(Microturbo TCS 14) was used to equip the Poyaud 520-6L S3 enginedescribed previously in conjunction with FIGS. 14-16 which develops amean effective pressure of 21 bar at 2,500 r.p.m., i.e., a power factorof 60 h.p. per liter of displacement in continuous operation.

The axial turbine is restricted by an expansion ratio of about 2.5.Thus, it is expected that turboblowers for high powered engines of theinvention will be equipped with a turbine having two axial wheels and anaxial compressor followed by a centrifugal compressor. However, untilsuch time as such turbines are available, supercharging of some enginespursuant to the invention may be accomplished with to series-mountedturboblowers.

It is also to be understood that the supercharger air cooler must beredimensioned for the new conditions at the compressor outlet. Watercooling may be retained when the cold source is natural water (maritimepropulsion) or when it is at some distance from the engine. In thiscase, present technology may be retained. Air cooling is alwayspreferable when the cold source is the atmosphere (railway traction).For this air/air exchanger brazed steel should be selected rather thanlight alloy which loses its mechanical properties above 300° C.

It thus will now be appreciated that the method and system of theinvention lies at the crossroads of two rapidly evolvingtechnologies--turbomachines and heat exchangers-- and of thewell-established conventional technology of the diesel engine whichbenefits from an enormous accumulated experimental knowledge. Pursuantto the present invention, and contrary to a widespread idea, the gasturbine is not a competitor of the diesel engine but is its opportunecomplement. After a very spectacular development in the aeronauticalfield, the gas turbine is today much closer to its asymptoticperformances than the diesel engine. In fact, if the laws of similaritywhich govern the dimensions of diesel engines are taken into account itwill be observed that by doubling the average pressure for a givenrotary speed, supercharging pursuant to the invention reduces the weightand bulk by a factor 2√2=2.83. This allows a weight to power ratio of1.4 Kg. per horsepower to be anticipated in the near future forindustrial engines with a power of less than 1500 h.p., decreasing evento much less than 1 Kg. per horsepower for light engines. Moreover,apart from this gain in performance, which places the diesel engine atthe same level as the industrial turbine with a heat recovering system,supercharging pursuant to the invention assists in reducing the priceper kilowatt installed and this is always an advantage.

From the foregoing description and drawings, it will now be understoodthat the various species of the invention disclosed herein embrace ageneric concept or concepts with respect to the manner in whichsupercharging is performed according to the method of the invention andaccomplished by the structure of the invention. The bypass, whether itbe a simple wide open pipe of fixed dimensions or geometry, as shown forexample in FIG. 1, or a bypass or regulated variable cross section, asshown for example in FIG. 9, is selected to have, or is provided with,air flow characteristics matched to the output of compressor 2 or 103such that any appreciable pressure difference developed in the airpassing through the bypass is generally independent of the ratio of theflow rate of the air traversing the bypass to the total air flowdelivered from the compressor. Thus, in the case of the constantly wideopen bypass of fixed dimensions shown in FIG. 1, the bypass, asindicated previously herein, is made large enough in flow capacity so asto permit the whole of the flow of air delivered by the compressor topass without appreciable pressure drop regardless of the speed of theengine, a condition which necessarily follows from the previouslydescribed operation of the turbo-compressor 2-4 at the aforementionedminimum threshold speed both while the engine 1 is running and while theengine 1 is stationary (and thus acting as a closed valve in the branchof the compressor-turbine flow path which passes through the combustionchamber system of engine 1).

In the event that a significant pressure drop is to be created in thebypass, the pressure drop must be controlled in accordance with thepresent invention by the previously described throttle means 108 or itsequivalent so that such pressure drop is an increasing function of thepressure existing at the compressor outlet to thereby maintain suchpressure drop independent of the ratio of the flow rate of the airtraversing the bypass to the total air flow delivered from thecompressor, or, stated another way, to the ratio of the flow rate of airtraversing the bypass to the flow rate of air traversing the enginecombustion chamber. Preferably, this is accomplished by suitablymatching the turbo-compressor to the engine combustion chamber systemand to the bypass so that the compressor always operates near enough toits surge line to optimize the efficiency of the compressor. Thus, for agiven pressure ratio of the compressor (i.e., for a given power outputof the engine), the air flow delivery of the compressor remainsgenerally constant regardless of the rotary speed of the engine (i.e.,fast, slow, stationary or accelerating or decelerating), therebyenabling operation of the compressor such that a plot of the pressureratio of the compressor versus the air flow delivery of the compressorwill lie generally along a line which extends generally parallel to thecharacteristic surge line of such compressor as shown in FIG. 23.

To summarize the above in slightly different terms, whether consideredfrom the standpoint of proper selection of a permanently wide openbypass pipe of fixed dimensions with respect to a properly matchedturbo-compressor and engine, which may be any of the arrangements shownin FIGS. 1-8 (which are particularly useful for nonscavenged engines),or a bypass having means for regulating the flow cross section such asshown in FIGS. 9-13 (which are useful for both scavenged andnonscavenged engines), the invention contemplates the correlation of theair flow permeability of the bypass with that of the engine combustionchamber system so that the combined air flow permeability or capacity ofthe parallel combination of air flow paths provided by the engine and bythe bypass means remains substantially constant regardless of the speed,or the rate of change of speed, of the engine in order to obtain theefficiencies and advantages described hereinabove.

Although the working examples particularly disclosed in the foregoingdescription deal with a method of supercharging an engine of thecompression ignition (diesel) type, it will be evident from theforegoing disclosure to those skilled in the art that this generalconcept of very high supercharging of internal combustion engines,featuring the "high yield mode of operation of the turbo-compressor,"which consists of controlling, by way of the above-described "bypasssystem," the difference between the compressor outlet pressure and theturbine inlet pressure, is also applicable to any type of superchargedinternal combustion engine of the expansible combustible chamber type.

I claim:
 1. A method of operating an assembly of a supercharged dieselinternal combustion engine and a turbine-compressor supercharger inwhich the turbine drives the compressor, said engine having an airintake manifold communicating with the compressor output and an exhaustmanifold communicating with the turbine and having a compression ratioless than 12, a gas flow path being provided from the compressor to theturbine, a combustion chamber in the gas flow path, said methodcomprising the steps of:(a) starting the turbine-compressor superchargerin rotation by separate starting means so as to initiate gas flowthrough the gas flow path to the turbine, (b) igniting fuel in thecombustion chamber to heat gases flowing through said gas flow path andentering the turbine so as to bring the turbine-compressor superchargerinto self-maintaining operation at or above a minimum threshold speedsufficient for the compressor to deliver air at temperature and pressureconditions high enough to provide self-ignition of the air fuel mixturein the engine at the end of its compression strokes when the engine isjust about to start, (c) starting said engine with said air from thecompressor at said temperature and pressure conditions entering theengine intake manifold, (d) thereafter, while operating said engine andsupercharger assembly after start-up, maintaining the speed of theturbine-compressor supercharger at or above said minimum threshold speedwhich provides temperature and pressure conditions sufficient to enablesaid self-ignition at the end of the compression strokes when the enginehas begun to operate, said minimum threshold speed having a fixed valuefor an engine having a given compression ratio, said minimum thresholdspeed being lower for an engine having a higher compression ratio andbeing higher for an engine having a lower compression, ratio and (e)concurrently with each of said steps (a) through (d) maintaining saidgas flow path constantly open.
 2. A method according to claim 1 whereinsaid minimum threshold speed is obtained by maintaining a sufficientflow of fuel to the combustion chamber, at least on starting.
 3. Amethod according to claim 2 wherein said flow of fuel is regulated. 4.The method of claim 1 comprising the further steps of:(e) operatingturbine-compressor such that the temperature of gases entering theturbine is sufficient to insure that the compressor outlet pressure ishigher than the turbine inlet pressure; and (f) controlling suchpressure difference between the compressor outlet and the turbine inletby the way of passageway means provided in said gas flow path such thatsaid pressure difference is generally independent of the ratio of theflow of the air traversing said gas flow path to the total air flowdelivered from said compressor whereby for a given pressure ratio of thecompressor the air flow delivery of said compressor remains generallyconstant regardless of the rotary speed of said engine, thereby enablingoperation of said compressor such that a plot of the pressure ratioversus the air flow delivery of said compressor is generally along aline generally parallel to its surge line.
 5. The method of claim 4wherein said passageway means has a flow controlling cross section witha given minimum dimension capable of permitting the whole of the flow ofair delivered by the compressor to pass through said gas flow pathwithout appreciable pressure drop, and wherein step (f) is performed bymaintaining said given dimension at a fixed value.
 6. The method setforth in claim 5 wherein said turbine-compressor is matched to saidengine and to said gas flow path such that the compressor alwaysoperates sufficiently near its surge line to optimize the efficiency ofsaid compressor.
 7. The method set forth in claim 4 wherein saidturbine-compressor is matched to said engine and to said gas flow pathsuch that the compressor always operates sufficiently near its surgeline to optimize the efficiency of said compressor.
 8. The method setforth in claim 4 wherein step (f) thereof is performed by controllingthe flow characteristics of said passageway means with a throttle valvemeans so as to cause a pressure drop in the air passing through the gasflow path which is an increasing function of the pressure existing atthe compressor outlet.
 9. The method set forth in claim 8 wherein saidflow characteristics are controlled such that said pressure drop is anincreasing linear function of the pressure existing at the compressoroutlet.
 10. The method set forth in claim 8 wherein saidturbine-compressor is matched to said engine and to said gas flow pathsuch that the compressor always operates sufficiently near its surgeline to optimize the efficiency of said compressor.
 11. A method ofoperating an assembly of a supercharged engine and an associatedsupercharger wherein the engine has an internal combustion chambersystem of the self-ignition type and wherein the supercharger comprisesat least one compressor and at least one turbine driving the compressorsuch that the supercharger drive is mechanically independent of theengine, said engine combustion chamber system communicating with thecompressor output and the turbine inlet and having a given compressionratio less than that necessary to achieve self-ignition with airadmitted at ambient temperature and pressure conditions, and theassembly also has a gas flow path between the compressor and the turbinein bypass relation to said engine combustion chamber system, said methodcomprising the steps of:(a) prior to starting the engine, starting theturbine-compressor supercharger in rotation so as to initiate gas flowthrough the gas bypass path to the turbine with sufficient input energyso that air is expelled from the compressor at correlated minimumtemperature and pressure conditions high enough such that, when said airis to be compressed by said engine in said combustion chamber systemfrom a point at or near the beginning of the intake phase of the cycleof said engine to a point at or near the end of the compression phase ofsaid cycle, self-ignition of the air-fuel mixture in the combustionchamber system of the engine will be produced when the engine is crankedto start; (b) then starting said engine while delivering said air fromthe compressor at said correlated minimum temperature and pressureconditions into the engine combustion chamber system; (c) thereaftermaintaining the speed of the turbine-compressor supercharger at or abovea minimum threshold speed which provides to said engine combustionchamber system air at said correlated minimum temperature and pressureconditions while the engine is operating under its own power, saidminimum threshold speed having a fixed value for said engine when allother conditions affecting self-ignition are constant and which isinversely dependent upon said ratio, and (d) concurrently with each ofsaid steps (a) through (c) maintaining said gas flow path constantlyopen.